EP3706928B1 - Rolling stand having a seal to prevent lubricant escaping - Google Patents

Description

The invention relates to a roll stand with a seal for sealing a lubricant space against leakage of lubricant.

Ring seals with cavities to achieve a dynamic sealing effect depending on the pressure conditions in the ring or lubricant gap of a bearing are e.g. B. known from the U.S. 5,169,159 or the EP 3 098 486 A1 . Roll stands with seals, as they are the basis of the present invention, are basically known in the prior art, for. B. from the German Offenlegungsschriften DE 10 2013 224 117 A1 or DE 10 2015 209 637 A1 . Both laid-open documents each disclose a roll stand with at least one roll for rolling, in particular, metallic rolling stock. The roll has two roll journals and a roll barrel. The roll stand has chocks - here for example each with a bearing bush - for the rotatable mounting of the roll in the roll stand. The chocks or the bearing bushes each span a receiving opening for receiving one of the roll journals, optionally with a drawn-on journal bushing. An annular gap for receiving a lubricant is formed between the roller and the chock. Both at its end on the ball side and at its end remote from the ball, the annular gap is sealed by an annular seal detachably attached to the bearing bush. The ring seal is designed in such a way that it prevents or reduces a lateral outflow of the lubricant from the annular gap in a predetermined circumferential angle range. For this purpose, sealing systems known from the prior art, such as. B. used radial shaft seals or labyrinth seals. The preamble of claim 1 is based on DE 10 2015 209 637 A1 .

However, all sealing systems used or tried out turned out to be unsuitable for the following reasons: The preload with which the Ring seal is pressed in the radial direction on the roll journal or the journal bushing, is structurally determined when designing the roll stand. The preload does not adapt to the pressure of the lubricant which varies in time and place in the annular gap. The result is that if the predetermined preload is too small, the annular gap is only inadequately sealed; ie there are leaks. In the opposite case, ie if the predetermined preload is too great, there is increased friction between the seal and the roll journal or the journal bush, as a result of which excessive wear of the seal occurs. This increased wear can quickly lead to the destruction of the seal. In addition, if the preload is too high, the seal can be destroyed by extrusion into an annular gap with low or atmospheric pressure. The seal typically does not extrude into the annular gap between the roll neck and the chock because there is a very high counterpressure there. Instead, there is a risk that the seal will extrude in the opposite direction because there is no counter pressure there.

Also known are ring seals for sealing ring gaps in roll stands, which have a circumferential groove open towards the ring gap on their side faces facing the ring gap.

The pressure in this circumferential groove - and thus the sealing effect of the ring seal - is, however, the same in all circumferential angular areas because the pressure is equalized over the circumference. This known ring seal with a circumferential groove is therefore not suitable for taking into account different pressure ratios acting in individual circumferential angular areas by means of differently strong sealing effects.

The invention is based on the object of developing a seal in a known roll stand to the effect that the pressing force with which the seal against a surface in the roll stand, for example the surface of the Journal bushing or the roll journal is pressed, is suitably adapted or set to the local pressure conditions in the annular gap to be sealed.

This object is achieved by the subject matter of claim 1.

In the context of the present invention, the term “bearing” means in particular - but not only - oil film bearings.

In the present invention, a distinction must be made between two types of annular gaps, also referred to below as lubricant spaces:
A first annular gap or lubricant space is part of the actual bearing, in particular an oil film bearing; it is therefore also referred to below as the annular gap or lubricant space of the bearing. It is formed between the bearing bush and the trunnion bush. If the roll stand is designed without a bearing bush and journal bushing, it is located between the chock and the roll journal. The lubricant in this first annular gap bears the entire load during the rolling operation, at least in places.

A second annular gap or lubricant space is not part of the actual bearing, but adjoins it. It is also referred to below as the annular gap or lubricant space under the seal. It is located between the running surface of the seal and the opposite journal bushing or the opposite roll journal. If the first annular gap is sealed on both sides by a seal, there are two second annular gaps under the seals per roll in a roll stand.

Both annular gaps or lubricant spaces are in fluid-conducting connection with one another. However, the radial height and thus the volume of the first annular gap are significantly greater than in the case of the second annular gap. That's why it stands Lubricant in the first annular gap during operation of the roll stand under a much greater pressure than in the second annular gap. However, the pressure in the first annular gap is typically distributed differently over the circumference.

As a result of the stressed fluid-conducting connection, the cavities in the seal are each individually connected in a fluid-conducting manner to the pressurized lubricant space of the bearing. As a result, when the seal is used, the cavities fill with the lubricant and, moreover, the cavities are also exposed to the same pressure, which varies in the circumferential direction and over time, as the lubricant in the lubricant chamber of the bearing. Due to the elasticity of the material from which the seal is made, the cavities expand to a greater or lesser extent depending on the strength of the pressurization. Due to the isotropic properties of the sealing material, the expansion of the cavities leads to an increase in the volume of the seal and thus automatically to an increase in the pressure with which the seal acts on a surface to be sealed, in particular a journal bushing or a roll journal. With variable pressure ratios, the volume of the seal and the pressing force also vary accordingly. In other words, the claimed design of the seal advantageously ensures that the pressure force with the pressure conditions in the lubricant chamber of the bearing is distributed over the circumference and varies over time, or that the pressure force automatically adapts to the pressure conditions in the lubricant chamber in a suitable manner. The adaptation of the pressing force to the different circumferential pressure conditions in the lubricant space of the bearing is achieved in the seal according to the invention in particular that no circumferential groove, but rather a plurality of cavities formed separately from one another in the circumferential direction are provided, which are in different circumferential areas with the lubricant space Bearing are in fluid communication.

For the seal according to the invention to function, it is important that, as said, the lubricant and the pressure to which it is exposed can get into the cavities; that is, there must be a fluid-conducting connection between the cavities and the pressurized lubricant space of the bearing. This fluid-conducting connection can, for example, either be implemented in that the cavities are formed as a recess on the surface of the seal and are open to the lubricant chamber of the bearing and / or that the cavities on the surface or inside the seal are fed into the lubricant chamber of the Open into the camp. In both variants, the cavities are in fluid or pressure-conducting connection with the lubricant in the lubricant space of the bearing.

By profiling the running surface of the seal, the properties of the lubricating film in the second annular gap between the seal and the contact surface of the journal bushing or of the roll journal can be adjusted.

The individual cavities in the seal can be designed differently in shape and size; they are preferably designed differently in groups. The strength of the pressing force can also be controlled with the size and shape of the cavities. The larger the cavities, the greater the achievable increase in volume of the seal and, consequently, the achievable pressing force and vice versa.

According to a further exemplary embodiment, the individual separate cavities in the seal are preferably evenly distributed over the length or the circumference of the seal. This offers the advantage that the seal has the same properties in every circumferential angular position and thus - assuming the same pressure - the same sealing effect can be achieved.

For the use of the seal in roll stands for sealing annular gaps, ie annular lubricant gaps of bearings, the training is the seal makes sense as a ring seal. When designed as a ring seal, the running or sealing surface of the seal is typically formed on the inside of the ring seal, ie facing the center or midpoint of the ring seal. This ensures that, particularly when the seal is used on roll necks, the sealing surface faces the outer surface of the roll neck.

The advantages of the claimed roll stand essentially correspond to the advantages mentioned above with reference to the claimed seal. The surface of the roll neck now forms the contact surface against which the seal is pressed with its running surface. As described above, due to the special structural design of the seal, the pressing force is now automatically adapted to the possibly very high pressure level in the annular gap between the roll neck and the chock or in the annular gap between the roll neck and the running surface of the seal.

This said variation of the pressing force as a function of the pressure conditions in the annular gap can be superimposed on a preset preload with which the running surface of the seal presses on a contact surface. The total radial pressure force realized in this way from the superimposition of preload and variable pressure force must be balanced or adjusted so that it is not too large on the one hand, but also not too small on the other. The total pressing force must not become too great because a lubricating film must be maintained between the running surface of the seal and the contact surface of the rotating roll neck in order to prevent solid body friction between the seal and the rotating roll neck. The solid body friction would have the consequence that the seal would wear and thereby become unusable or destroyed in the long run. On the other hand, the total sealing force must not be too small, so that the thickness of the lubricating film is significantly less than the thickness / height of the first annular gap between the roll neck and the chock. Only if the thickness of the lubricating film in the second annular gap is significantly less than the thickness of the lubricating film in the first annular gap, the seal can also bring about the desired - not absolute, but extensive - sealing effect.

In the case of hydrodynamic oil film bearings in particular, a complete seal is typically not desired, but only a seal in a specific circumferential angle range, namely where the smallest lubricant film thickness occurs. Therefore, the seal does not necessarily have to be designed as a full circumferential ring seal; rather, z. B. for the said application, a sealing strip or ring segment of limited length in said circumferential angle range.

In the case of the claimed roll stand, a groove open towards the roll neck is formed on the end of the roll barrel side and / or on the end of the chock or the bearing bush remote from the roll barrel. The groove is used to accommodate the ring seal.

For example, the axially outside of the groove, i. H. the outside of the groove facing away from the bearing bushing can be formed by a perforated disk that can be releasably connected to the chock or the bearing bushing, for example a screwable perforated disk. Because the width of the ring seal in the unloaded state is greater than the width of the groove in the axial direction, the pretensioning force with which the seal acts on the roll neck in the radial direction can be adjusted by screwing on the perforated disk. The reason for this is that the elastic seal is isotropic when it is clamped in the axial direction in the groove. H. also expands in the radial direction. Depending on the pressure exerted by the screw connection in the axial direction, said pretensioning force can also be set in the desired manner.

To compensate for manufacturing tolerances in the seal and to ensure that the seal is at least slightly in its axial width is greater than the width of the annular groove, it is advantageous if the annular seal has at least one elevation on its front side facing the chock and / or on its front side facing away from the chock.

As already stated above, the recesses or cavities open directly into those surface areas of the seal which delimit the pressurized lubricant space of the bearing. Alternatively or additionally, however, further recesses can also be formed in those areas of the surface of the seal which, when the seal is clamped in the groove, are pressed against the walls or the bottom of the groove and are thus sealed by being pressed on. These further recesses or cavities are then preferably in fluid-conducting connection with the lubricant chamber of the bearing via the supply channels.

On its end face facing the ring seal, the chock or the bearing bush can preferably have pins protruding in the axial direction. These pins are arranged in such a way that they engage in said recesses on the surface of the seal. The pins advantageously serve, on the one hand, as an anti-twist device for the seal, in particular during the rolling operation, and, on the other hand, to limit the deformation of the seal. Another possibility is the radial arrangement of the pins with the same functionality.

The volume of the recesses can be greater than the volume of the pins which protrude into the recesses when the seal is installed. This embodiment offers the advantage that, despite the pins protruding into the recesses, a remaining cavity remains which, if it is connected to the annular gap via a feed channel, can function as a cavity within the meaning of the present invention.

A design of the inner diameter of the ring seal larger than the outer diameter of the roll neck - optionally with a drawn-on journal bushing - at the axial height of the ring seal offers the advantage that the bearing formed by the chock with the roll neck stored therein can be operated as an oil film bearing. The prerequisite for this is that the ring seal reduces the annular gap between the running surface and the contact surface only in a limited circumferential angle range, where the minimum lubricant film density prevails, down to the said lubricant film, while the seal does not have to achieve any significant sealing effect in the remaining circumferential angle range. In the area of the minimum lubricant film thickness, the entire pressing force is maximized and thus the annular gap is reduced down to the lubricant film. In the rest of the circumferential angular range, the pressing force is negligibly small and the lubricant can escape from the annular gap there in the axial direction due to the overdimensioning of the annular seal. This is what is intended for operation as an oil film bearing.

Further advantageous configurations of the seal and of the roll stand with the seal are the subject of the dependent claims.

Figur 1

die erfindungsgemäße Dichtung;

Figur 2

ein Walzgerüst aus dem Stand der Technik;

Figur 3

die Lagerung einer Walze in einem Einbaustück;

Figur 4

die Ausbildung und Anordnung der Dichtung in der Lagerung nach Figur 3 in einer vergrößerten Ansicht; und

Figur 5

einen Querschnitt durch den von einer Lagerbuchse oder einem Einbaustück aufgespannten Aufnahmeraum mit eingesetztem Walzenzapfen

zeigt.The description is accompanied by 5 figures, where

Figure 1

the seal according to the invention;

Figure 2

a roll stand from the prior art;

Figure 3

the storage of a roller in a chock;

Figure 4

the design and arrangement of the seal in the storage Figure 3 in an enlarged view; and

Figure 5

a cross-section through the receiving space spanned by a bearing bush or a chock with inserted roll journal

shows.

The invention is described in detail below in the form of exemplary embodiments with reference to the figures mentioned. In all figures, the same technical elements are denoted by the same reference symbols.

Figure 1 shows the seal 100 designed according to the invention. It can be used to seal a lubricant space (in Figure 1 not shown) serve against leakage of lubricant. It is made at least partially from an elastic material and has a plurality of arbitrarily shaped cavities 110 in its interior. The cavities 110 are in Figure 1 designed merely by way of example as cylindrical recesses 110 which are open towards the surface of the seal. Alternatively, the cavities 110 can also be formed completely in the interior of the seal; they are then connected to the surface of the seal in a fluid-conducting manner via supply channels 120. The supply channels serve to supply the lubricant from the lubricant space 300 of a bearing, in particular an oil film bearing, into the respective cavities 110.

In Figure 1 the underside of the seal 100 forms a running surface 112 with which the seal against a contact surface of a typically moving object, e.g. B. a roll neck is pressed.

At least largely, the recesses or the supply channels 120 open into a first lubricant chamber 300 of the bearing that is to be sealed. This ensures that the lubricant and the pressure are transmitted from the first annular gap 300 into the cavities. It then turns into the Cavities always enter the possibly varying pressure from the lubricant chamber. With regard to the purpose associated therewith, reference is made to the above statements in the general part of the description.

The reference numeral 225 and the hatching indicate wall areas of a groove into which the seal can typically be inserted. These wall areas then largely cover areas of the surface of the seal. Only the feed channels and / or recesses in the uncovered areas of the surface of the seal are in fluid-conducting connection with the first lubricant chamber 300; see also Figure 4 .

As in Figure 1 As can be seen, the recesses 110 can also be opened to other surface sections of the seal than to the lubricant chamber 300. This is particularly advantageous for these recesses to interact with the pins on the chock, which will be described below.

Figure 1 FIG. 12 also shows, by way of example, a first group of cutouts 110 and a second group of cutouts 110, the volume of the cutouts in the first group being greater than the volume of the cutouts in the second group. With the same applied pressure, the different volumes cause a different expansion of the seal and thus possibly a different proportion of a sealing force exerted by the seal.

The seal in Figure 1 has a width a; this width also corresponds, for example, to the width of the running surface 112. It can also be seen that the seal has, for example, a rectangular cross section. The separate cavities 110 or recesses 110 are preferably evenly distributed over the length or the circumference of the seal. This has the advantage that the seal 100 therefore also dies at each point or each length section has the same properties. In order to seal cylinders, such as a roll neck 212, the seal 100 can be designed in the form of an annular ring as a ring seal; please refer Figure 5 . The running surface 112 is then designed to face the center or the midpoint of the ring seal or, in other words, the surface or contact surface of the cylinder; please refer Figure 5 .

Figure 2 shows a roll stand according to the prior art, as is also the basis of the present invention. The roll stand 200 has at least one, here by way of example four rolls 210, each with two roll journals 212 and one roll barrel 214 each. In particular, those in FIG Figure 2 The two middle work rolls shown are used for rolling rolling stock. The rollers 210 are each rotatably mounted with their roller journals 212 in a chock 220, also called a bearing housing.

Figure 3 shows this storage in a longitudinal section in detail. The roller 210 can be seen with its roller journal 212 and its roller barrel 214. A journal bushing 216 is pulled onto the roller journal. The roll journal with the journal bush is mounted in a receiving opening which is spanned by a bearing bush 222. The bearing bush 222 is arranged non-rotatably in the chock 220. An annular gap 300 is formed between the non-rotatably arranged bearing bush 222 and the journal bush 216 rotating with the roll neck 212 and is filled with lubricant 320 during operation of the roll stand. The lubricant is then under a high pressure, typically a few 100 bar, in the annular gap. In Figure 3 For example, the annular gap 300 is sealed by the ring seal 100 according to the invention both at its end on the roller barrel side and at its end remote from the roller barrel. The ring seal 100 does not have to be designed according to the invention over its entire circumference; in principle, only a section of the ring seal can be designed accordingly.

In Figure 3 it can also be seen that the bearing bushing 222 has, at its end on the side of the roll barrel and at its end remote from the roll barrel, a groove 230 which is open towards the roll journal 212 and into which the ring seal 100 is inserted. The outsides of the two grooves 230 are in the in Figure 3 The example shown is not formed by the bearing bush 222. Rather, the outer sides of the grooves are each formed there by perforated disks 240 which are screwed onto the bearing bush 222 with screws 245. Because the width a of the ring seal 100 in the unloaded state and, if necessary, taking into account the elevations 130, see Figure 1 are deliberately designed slightly larger in the axial direction R according to the invention than the width A of the groove structurally predetermined by the bearing bush 222, it is possible that by tightening the screws 245 the axial force with which the ring seals 100 are squeezed in the groove is variable can be adjusted. Due to the isotropic behavior of the material of the seal 100, the axial squeezing or upsetting causes not only a reduction in the width of the ring seal, but also an expansion of the ring seal in the radial direction. A variation in the axial clamping force therefore also automatically causes a variation in the preload or the radial pressing force with which the running surface 112 of the seal 100 is pressed against the opposite contact surface of the trunnion bushing 216.

Figure 4 shows the installation of the seal 100 in the bearing bush 222 again in detail.

The reference numeral 225 denotes the wall areas of the groove in the chock or the bearing bushing, against which the seal 100 is pressed when it is installed in the groove. In other words, the seal and the cavities on the surface of the seal are optionally covered and sealed by these wall areas. Cavities or feed channels arranged only radially further inward open into the first annular gap 300. It can also be seen that the thickness of the lubricating film 330 in the second annular gap 140 between the running surface 112 of the seal 100 and the opposite contact surface 218 of the journal bushing 216 is significantly less than the thickness of the annular gap 300 largely protrudes in the radial direction into the annular gap 300 originally present there. In Figure 4 the proportions are shown exaggerated. The ring seal actually presses on the contact surface 218 of the trunnion bushing 216 due to the radial pressing force FR. Due to the aforementioned high pressure ratios, however, the said lubricating film 330 with a thickness of only a few μm forms between the running surface 112 and the contact surface 218. The pressure in the first annular gap 300 is significantly greater than in the second annular gap 140.

It can also be seen that the ring seal 100 is inserted into the groove 230 in such a way that its recesses 110 engage with pins 228 which extend from the end face of the bearing bush 222 delimiting the groove 230, preferably in the axial direction R. In the case of radially arranged cavities, the pins extend perpendicular to the axial direction. The remaining cavity 116 is sealed off from the bearing bush 222 due to the axial or radial pressing force with which the ring seal 100 is pressed into the groove with the aid of the perforated disk 240; it therefore functions as a cavity 110 in the sense of the invention, which opens into the lubricant chamber 300 of the (oil film) bearing via a feed channel 120. The effect of the variation, in particular the enlargement, of the total radial force as a function of the pressure conditions in the region of the lubricating film 330 has been described in detail above.

Figure 5 has already been briefly described in the introduction. It shows a cross section through the receiving space spanned by the bearing bush for receiving the roll neck 212. As seen from a synopsis of the Figures 4 and 5 can be seen, the receiving space is not only provided by the bearing bush 222, but in particular also limited in the radial direction by the ring seal 100, which typically protrudes further inward. Inside the receiving space, the roll journal 212 rotates, if necessary with the journal bushing 216 drawn on. For operation of the roll bearing as a hydrodynamic oil film bearing, the inner diameter dD of the ring seal 100 is greater than the outer diameter DZ of the roll journal 212 at the axial height of the ring seal, optionally with the journal bushing 216 drawn open. During operation as a hydrodynamic oil film bearing, the ring seal 100 then presses itself so closely against the surface or contact surface 218 of the roll neck that only the Lubricant film 330 forms. In the remaining circumferential angular range, the running surface 112 of the ring seal 100 no longer rests on the contact surface 218 of the roll neck 212; rather, the distance between these two surfaces is greater than the thickness of the lubricating film in the region of the maximum rolling force; this applies in particular to the oversizing of the ring seal 100. The oversizing and the resulting larger gap between the running surface 112 and the contact surface 218 advantageously enable an axial drainage of lubricant in the circumferential angle area outside the area of the maximum rolling force. The outer diameter DD of the ring seal 100 typically corresponds to the inner diameter of the bottom of the groove 230.

List of reference symbols

100

poetry

110

Cavity or recess

111

Tread profiling

112

Tread

116

remaining cavity

120

Feed channel

130

Cant

140

Annular gap or lubricant space under the seal

200

Roll stand

210

roller

212

Roll neck

214

Roll barrel

216

Trunnion

218

Contact surface

220

Chock

222

Bearing bush

225

Wall areas of the groove

228

pen

230

Groove

240

Perforated disc

245

screw

300

Annular gap or lubricant space of the bearing

320

lubricant

330

Lubricating film

a

Width of the ring seal in the axial direction

A.

Width of the groove in the axial direction

DD

Outer diameter of the ring seal

Double room

Outer diameter of the roll neck, if necessary with neck bush

dD

Inner diameter of the ring seal

R.

axial direction

FR

radial pressure on the seal

FWmax

maximum rolling force

Claims (11)

1. Roll stand (200) comprising:

   at least one roll (210) with two roll journals (212) and a roll barrel (214) for the rolling of a rolling material;

   at least one chock (220), optionally with a bearing bush (222), for rotatable mounting of the roll (210) in the roll stand (200), wherein the chock (220) or the bearing bush spans a receiving opening for reception of one of the roll journals (212), optionally with a drawn-on journal bush (216), wherein the inner diameter of the receiving opening is formed to be larger by comparison with the outer diameter of the roll journal, optionally with drawn-on journal bush, in such a way that an annular gap (300) for receiving a lubricant (320) is formed between the chock or the bearing bush and the roll journal or the journal bush; and

   a seal, which is arranged at the end at the chock (220) or bearing bush (222) at the roll barrel side and/or at the end of the chock (220) or bearing bush (222) remote from the roll barrel to be secure against rotation relative to the rotatable roll, for sealing the annular gap - which represents a lubricant chamber - at least in a predetermined circumferential angular region;

   characterised in that

   the seal (100) is formed at least partly from a resilient material;

   the seal has at least two cavities (110) which are separate from one another in circumferential direction and which are open towards the lubricant chamber of the bearing in order to feed the lubricant (320) from the lubricant chamber (300), which is to be sealed, of the bearing to the cavities;

   a groove (230) which is open towards the roll journal (212) and into which the seal (100) is insertable is formed at the end of the chock (220) or bearing bush (222) at the roll barrel side and/or at the end of the chock (220) or bearing bush (222) remote from the roll barrel; and the axial width a of the seal (100) in unloaded state is greater than the width A of the groove (230) in axial direction R.
2. Roll stand according to claim 1, characterised in that the seal (100) is constructed as a segment of a ring and has a predetermined limited length L in which L < total circumference of the annular gap.
3. Roll stand according to claim 1, characterised in that the seal (100) is constructed in the form of an annular seal.
4. Roll stand (200) according to claim 3, characterised in that the outer diameter (DD) of the annular seal in unloaded state is substantially equal to the diameter of the groove (240) at the base thereof.
5. Roll stand (200) according to one of claims 3 and 4, characterised in that the inner diameter (dD) of the annular seal (100) is greater than the outer diameter (DZ) of the roll journal (212), optionally with drawn-on journal bush (216), at the axial height of the annular seal.
6. Roll stand (200) according to claim 1, characterised in that the axial outer side of the groove is formed by an apertured disc (240) which is detachably connectible, for example screw-connectible, with the chock or the bearing bush (222).
7. Roll stand (200) according to any one of claims 1 to 6, characterised in that the seal (100) has at least one elevation (130) at the end thereof remote from and/or facing the chock.
8. Roll stand (200) according to claims 1 to 7, characterised in that the seal (100) has a rectangular cross-section and the sealing surface of the seal faces the roll journal.
9. Roll stand (200) according to any one of claims 1 to 8, characterised in that the recesses (110) at the surface of the annular seal are also formed to be open towards the chock (220) or the bearing bush (222).
10. Roll stand (200) according to claim 9, characterised in that the chock (220) or the bearing bush (222) has, at the end face thereof facing the seal, pins (228), which project in axial or radial direction, for engagement in the recesses (110) at the surface of the seal (100).
11. Roll stand (200) according to claim 10, characterised in that the volume of the recesses (110) into which the pins project is greater than the volume of the pins (228) engaging in the recesses.

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